1. Field of the Invention
The present invention relates to a rotor blade damping structure for an axial-flow turbine, and more specifically to an improvement in the structure of rotor blades for an axial-flow turbine to reduce dynamic stresses and to obtain superior damping properties.
2. Description of the Prior Art
An axial-flow turbine is driven by fluid flowing between rotor blades arranged in the circumferential direction of a rotor so as to form an annular blade arrangement, and energy is transmitted from the fluid to a rotor shaft through the rotor blades. With the recent trend toward increases in the capacity of electric power plants, the volume of flow has increased more and more and the operating conditions (e.g., operating temperature and pressure) have become more and more severe, with the result that the various forces applied to the rotor blades have increased more and more. These forces inevitably cause various internal stresses such as centrifugal stress, thermal stress, bending stress, torsional stress, etc, in the turbine rotor blades, and sometimes generate violent vibration stresses in the rotor blades independently or in combination. Accordingly, it is an important problem to consider how to cope with blade vibration, that is, how to obtain a large dynamic stress reduction and superior damping properties.
One method of reducing the turbine blade dynamic stress is to link a plurality of adjacent turbine rotor blades together by use of a rigid link member. With this method, however, there is the problem that stress is often concentrated at the linkage or interconnection points between adjacent turbine rotor blades. In addition, a torsional stress is inevitably generated in the rigid link member due to the untwisting of the rotor blades during turbine rotation (by centrifugal force), and this problem must be solved. Further, in the type where holes are formed through the rotor blades to link the blades with wire, for instance, a problem arises in that stress readily concentrates around the holes and the holes undergo corrosion with the elapse of time with resultant accumulation of corroded compositions in the holes. On the other hand, under the present situation wherein turbine units become superannuated more and more in the electric power plants, when the above mentioned link members are used for the turbine rotor blades, the blades cannot be detached easily from the turbine, and there arises another problem in that it is difficult to inspect the quality of the rotor and blade dovetail attachment portions to check the remaining life time.
As another method of reducing dynamic stress of the turbine rotor blades, a snubber structure is also well known wherein a shroud is formed integrally with each blade at the top end thereof in such a way that the shrouds of adjacent blades are brought into contact with one another during turbine rotation. A typical example of this snubber structure will be described in further detail below with reference to FIG. 15.
In FIG. 15, blades 1 are assembled to a rotor 2. A shroud 3 is formed integrally with each blade 1 at the top end thereof. Adjacent shrouds 3 are brought into contact with each other during turbine rotation. These adjacent shrouds 3 are assembled so as to provide a minute gap therebetween (a snubber gap) at rest. During turbine rotation, however, the gap is eliminated by the phenomenon that the twisted blade 1 is untwisted by centrifugal force, and the two adjacent shrouds are brought into pressure contact with each other at the end surfaces thereof, and thus the blade vibration is reduced as a result of a vibration damping properties due to the pressure contact of the shrouds.
FIG. 16 is a view of the blades as seen from the blade top radially inward, in which the dashed lines represent the blades 1 when at rest and the solid lines represent the blades during rotation. As depicted in FIG. 16, the snubber gap existing between two adjacent shrouds 3 during the non-rotating condition is eliminated due to the untwisting of the blades caused by centrifugal force applied to each blade 1, so that the two adjacent shrouds 3 are brought into contact with each other.
FIG. 17 shows a blade 1 represented by a twisted plate for simplicity, in which the solid lines show the blade during rest and the dashed lines show the blade during rotation. That is, when the twisted plate 1 shown by the solid lines is pulled at both ends thereof in two opposite arrow directions A, the twisted plate shown by the solid lines is untwisted to the state shown by the dashed lines. In the same way as above, the blade 1 in FIG. 15 is pulled in the longitudinal direction A during rotation, so that the blade 1 is untwisted.
As described above, in the snubber structure, shrouds assembled so as to provide a minute gap between adjacent shrouds, can be brought into contact with one another by the utilization of the untwisting force of the twisted blades. And the blade dynamic stresses could be reduced by friction of contact.
FIGS. 18(a) and (b) show another example of the snubber structure, in which FIG. 18(a) shows a single blade 1 (dashed line) and a single shroud 3 as seen from the top end of the blade, and FIG. 18(b) shows a plurality of blades 1 (dashed lines) and a plurality of shrouds 3 in their assembled state. In FIG. 18(a), a contact surface 4 of the shroud 3 has an inclination angle .theta.1 with respect to the axial direction of the turbine, and the pitch l1 between the two side contact surfaces of the shroud 3 is set to a value slightly larger than a geometrical pitch calculated on the basis of the diameter of the shroud contact surface and the number of blades. On the other hand, in the assembled state shown in FIG. 18(b), the blades are twisted to provide a torsional angle .theta.2 between the blade root portion and the shroud 3 and the pitch between the two side contact surfaces of the shroud 3 is set to a geometrical pitch l2. Therefore, in assembled condition, a surface pressure can be generated between the contact surfaces of two adjacent shrouds 3 due to the untwisting force on the twisted blades, so that the vibration damping properties can be obtained. FIGS. 19(a) and (b) show still another example of the snubber structure, in which FIG. 19(a) shows partially assembled blades as seen from the rotor axial direction. In FIG. 19(a), a shroud 3a provided for the blade la is formed with two opposite tapered surfaces converging radially outward of the blade 1a, and a pair of shrouds 3b provided for fixed blades 1b adjacent to the blade 1a are formed each with two opposite tapered surfaces converging radially inward of the blade 1b. FIG. 19(b) shows a blade 1a as seen along the rotor circumferential direction. In FIG. 19(b), a dovetail attachment portion 6a of the blade 1a is fitted in a groove 5 formed in the circumferential surface of the rotor 2. Further, when assembled, a gap m is given between a dovetail load bearing surface 7a of the blade 1 and a grove load bearing surface 8a of the rotor 2. In other words, the blade 1a is previously assembled to be offset radially inward so that it can be shifted radially outward by centrifugal force generated by the blade 1a during rotation. Therefore, when the blade 1a is shifted radially outward during rotation, the shroud 3a of the blade 1a is brought into contact with both the shrouds 3b of the blades 1b, so that all the shrouds are coupled with each other to form a continuously coupling structure throughout the circumference of the rotor blades.
One of the features of the blades of the snubber structure with respect to vibration is that all the blades arranged on the circumferential surface of the rotor can be continuously coupled in one ring by the coupling structure. In more detail, in the case where a plurality of blades are linked via rigid linking members 9 as shown in FIG. 20(a), there inevitably exist vibration modes in which grouped blades vibrate in the same phase together. In particular, the vibration mode in tangential direction of the rotor as shown in FIG. 20(b) is a low order vibration mode, and such a tangential mode is low in frequency and has higher dynamic stresses. In the case where all the blades are coupled together throughout the circumference of the rotor, even if an external force is applied to the blades so as to excite this vibration mode, the vibration energies cancel each other within the continuously coupled blades, and therefore there exists the advantage that stress level of tangential mode vibration is reduced against an external force applied to the rotor blades.
In the prior art rotor blade structures, however, there exist various drawbacks as follows:
In the untwist type snubber blade structure shown in FIG. 15, the centrifugal force is small when the rotor rotational speed is low and therefore the untwisting of the blades is small. Consequently there is the problem that the contact surfaces of the shrouds are not brought into pressure contact with one another perfectly, and a large dynamic stress reduction and damping properties cannot be expected.
In particular, when the blade length is large, the blades are designed in such a way that the natural frequency does not match the harmonic frequencies of the rotor rotation speed at the rated rotation speed, because a large exciting force is applied at the resonance of blade natural frequency and harmonic frequency. However, whenever the turbine is started or stopped, it is unavoidable that the rotor natural frequency matches the harmonic frequencies of the rotor rotation speed. When the shrouds are not brought into contact with one another under these conditions, the blades vibrate violently and may be broken in the worst case.
On the other hand, in the case where the blade length is relatively short, the blades are twisted to a small degree, and the untwisting of the blades hardly occurs at the rated rotation speed. In this case, therefore, it is impossible to apply the untwist type snubber structure to the short length blades.
Ideal conditions of the blades are that the blades are always provided with the dynamic stress reduction and damping properties under all circumstances, including acceleration or deceleration or rotation at the rated rotation speed. To achieve the above-mentioned conditions, it is necessary to always keep the snubber gap zero, that is, that adjacent blades are always in contact with one another under any operating conditions.
For that reason, the snubber gap must be kept zero in the assembled state. However, where the contact surfaces of the shrouds are in light contact with each other in the assembled condition, the rotor and blades are both elongated outward in the radial direction by centrifugal force during rotation, so that the overall diameter of the shrouds increases and thereby a slight gap is inevitably produced between two adjacent shrouds. As a result, it becomes impossible to keep the shrouds in contact with each other.
Under the above-mentioned conditions wherein two adjacent contact surfaces of the shrouds are opposed to each other with a slight gap therebetween or in light contact with each other, there exists a possibility that the contact surfaces of the shrouds are damaged, when the shrouds collide against each other, and consequently the contact surfaces are subjected to wear, thus deteriorating the blade reliability.
On the other hand, a large vibration damping properties can be obtained in this snubber structure as long as the snubber contact surfaces are in tight contact with each other with certain pressure. And when the shrouds are stably connected to each other as continuously coupled blades, it is possible to expect an effective vibration damping properties.
In the prior art blades of twisted type as shown in FIGS. 18(a) and (b), the blades are assembled with a twist produced between the blade root portion and the shroud, so that an initial surface pressure can be generated between the snubber contact surfaces in assembled condition due to elasticity of the airfoil portion. However, the torsional rigidity of the blade is extremely high in general, so that when a required torsional deformation is given to the blade an excessive internal stress is inevitably generated in the blade and the dovetail attachment portion. In particular, in the dovetail attachment portion (at which the blade is fixed to the rotor), the blade is brought into non-uniform (partial) contact with the rotor-side groove due to the torsional deformation of the blade-side dovetail portion, with the result that a high local stress is generated there. In addition, in the case of a blade of small length, in particular, the blade is slightly deformed by twisting, and therefore a larger local stress is generated in the blade dovetail portion. Further, small blades are usually used in high temperature and high pressure section of the turbine. Therefore, where the margin of the material strength is not sufficient, an increase in the additional torsional stress or the local stress is harmful on the blade reliability.
Further, in the twist type blade shown in FIGS. 18(a) and 18(b), during rotor assembly, each blade must be assembled to the rotor by pushing the blade against the adjacent blade with a strong force under the conditions that the blade is maintained twisted. Therefore, a special jig or stopper must be prepared, and consequently another problem arises in that the assembly work takes a long time.
On the other hand, in the blade formed with a wedge type shroud shown in FIG. 19(a) and (b), no blade torsional deformation is used, so that this snubber structure can be applied to a relatively short blade of high rigidity. Further, the shrouds of the adjacent blades can be brought into pressure contact with each other during the turbine rotation. However, there is a possibility that the offset shifted blades 1a will return again to their original positions after the turbine has stopped. Even if they do not return to their original positions naturally, when a small shock is applied to the blades, the offset blades tend to be easily returned to their original positions. Therefore, when the blades are shifted or moved at start and stop of the turbine, the above-mentioned blade movement causes abrasion in the contact surfaces between the shrouds and tends to damage the blade dovetail portions. This is not desirable from the viewpoint of rotor balance.
There is another possibility that even when the centrifugal force is applied the blade 1a cannot be shifted sufficiently due to the obstruction by the adjacent shrouds 3b of the fixed blades 1b, so that the turbine is rotated under the conditions that the load bearing surface 7a of the blade dovetail portion 6a of the blade 3a and the load bearing surface 8a of the rotor 2 are not brought into contact with each other. In this case, since all the centrifugal force of the blade 1a is applied to only the adjacent blades 1b, another problem arises in that an excessive local stress could be generated in the shroud, blade and blade dovetail portions of the adjacent blades 1b.
Further, in the prior art blades of this type, there exists another problem that no means is provided for adjusting the position of the blades in the rotor axial direction during assembly. In more detail, as shown in FIG. 19(b), there are gaps S2, S3 and S4 between the blade 1a and the rotor 2 in the axial direction of the rotor 2 in the fitting portion between the two. These gaps are inevitably produced due to machining tolerances of the blade 1a and the rotor 2, and it is impossible to reduce these gaps to zero. If these gaps are large, the snubber blade 1a will be shifted inclinedly relative to the axial direction according to the contact conditions between the wedge shaped contact surfaces of the shrouds. When the blade is shifted inclinedly relative to the axial direction, an imbalanced load will be applied to the load bearing surfaces of the dovetail portions and an excessive stress will inevitably be generated in the blade dovetail portions.